Microtube-strip heat exchanger

ABSTRACT

A new approach to the theory of heat exchanger optimization is presented which shows the advantages of using low Reynolds and Nusselt numbers and low flow velocities along with a novel design, the microtube-strip (MTS) counterflow heat exchanger. The MTS exchanger in the preferred embodiment consists of a number of small modules connected in parallel. Each module typically contains eight rows of one hundred tubes, each of 0.8 mm outside diameter and 0.16 m length. The tubes are metallurgically bonded via the diffusion welding technique to rectangular header tube strips at each end. Caps suitable for manifolding are welded over the ends. Cages are provided to cause the shell-side fluid to flow in counterflow fashion over substantially all of the tube length, and suitable manifolds are provided to connect the modules in parallel. This design results in the highest power densities of any known design for single phase exchangers. Although the MTS exchanger of the present invention is specifically optimized for applications not involving phase changes in the working fluid, the essential concepts and features of this invention can also be advantageously used in applications involving change of phase.

.Iadd.This application is a continuation of application Ser. No.243,048, filed on 9/9/88, now abandoned, which is a reissue ofapplication Ser. No. 700,125, filed on 2/11/85, U.S. Pat. No. 4,676,305..Iaddend.

BACKGROUND OF THE INVENTION

The field of this invention is heat exchangers, and more particularly,counterflow, modular, shell-and-tube-type exchangers for single phasefluids with no heat transfer augmentation means.

PRIOR ART

The result of four decades of industrial and commercial interest in heatexchangers has seen a proliferation of specialized devices andmanufacturing techniques that offer some advantages in specialapplications. The present invention is based on a radical departure fromconventional heat exchanger design guidelines in several distinct areas.As a result, the design differs in a number of ways, but the mostsignificant innovative feature is the most subtle and is not apparentwithout a detailed theoretical explanation. This most important featureis its size. This change represents such radical departures fromconventional practice in typical Nusselt and Reynolds numbers as to makereference to prior art of limited value. Nonetheless, for referencevalue and completeness, a brief synopsis of the prior art is presented.

Numerous examples of modular, counter-flow shell and tube exchangers canbe found in the patent literature, one of the earlier examples beingRossi's bi-directional flow design, U.S. Pat. No. 2,839,276, with itsadvantages of reduced thermal stresses. A more typical recent design isthat of Baumgaertner et al, U.S. Pat. No. 4,221,262, which offers someconstruction advantages over earlier designs due to the reducedcomplexity of its basic modules. Quite atypical and impractical, but ofrevelance on account of its general system appearance, is Giardina'sU.S. Pat. No. 4,253,516, with its huge box-car sized modules.

Jabsen et al in U.S. Pat. No. 4,289,196 and Culver in U.S. Pat. No.4,098,329 employ unique heading and manifolding systems in attempts toachieve higher power densities in modular systems. Cunningham et al giveattention to hot corrosive problems in U.S. Pat. No. 2,907,644.Lustenader recognizes the problem of axial conduction losses in U.S.Pat. No. 3,444,924, a problem obviously not understood by most heatexchanger design engineers.

Corbitt et al address the problem of vortex induced resonances in crossflow exchangers, U.S. Pat. No. 2,655,346, and solve it via the strategicpositioning of baffles. Scheidl uses a tube support grid to solve theseproblems in U.S. Pat. No. 3,941,188.

Bays, U.S. Pat. No. 2,537,024, and Malewicz, U.S. Pat. No. 3,452,814,give several examples of heat flow augmentation, which is easily shownto be of negative value in a gas-gas heat exchanger optimized accordingto the present invention.

Various well-known joining techniques include Cottone and Sapersteine'suse of special braze alloys, U.S. Pat. No. 4,274,483, Olsson andWilson's cold pressure welding, U.S. Pat. No. 4,237,971, Hardwick'sexplosive welding, U.S. Pat. No. 3,717,925, Brif and Brif's expandedtubes, U.S. Pat. No. 4,239,713, and the related technique of Yoshitomiet al, U.S. Pat. No. 4,142,581. More closely related to the diffusiontechnique of the present invention is the press-fit method of Nonnenmannet al, U.S. Pat. No. 4,159,741, and the compression method of Takayasu,U.S. Pat. No. 3,922,768. However, these techniques as described fallshort of producing a high integrity metallurgical bond. Lord's U.S. Pat.No. 4,528,733 describes a joining technique suitable for applications inwhich the header is made of a material which undergoes a phase changethat is accompanied by an abrupt change in dimension. Mattioli et al,U.S. Pat. No. 3,849,854, describe a method of effecting diffusion weldsvia induction heating followed by electromagnetic compression that issuitable for large, accessible joints. Woods, U.S. Pat. No. 2,298,996,describes a method of hard brazing aluminum and copper alloy 6 mm tubes,extending beyond their rectangular headers and expanding into apolygonal shape so as to reduce tube side pumping losses in turbulentflow applications, into rectangular headers, while Troy, U.S. Pat. No.3,782,457, describes the use of 2 mm tubes in an annular header withheat transfer augmentation.

Frei's U.S. Pat. No. 4,295,522, employing glass tubes and siliconecasting resins, shows a striking resemblance from a non-scaledperspective between his basic tube assembly modules and the presentinvention. Furthermore, the tube sizes employed therein also showprogressive design traits, being about 6 mm in diameter rather than thecustomary 1.5 cm to 2.5 cm employed in all other above referencedpatents. However, Frei's design, aside from temperature and pressurelimitations imposed by the choice of materials, suffers from theinefficiencies inherent in a cross-flow design, as necessitated by hismanifolding scheme.

The use of small diameter tubes-O.D. less than about 3 mm-has beenpredominantly limited to two-phase cross-flow systems. Early examplesmay be found in aircraft oil-coolers such as that by Anderson, U.S. Pat.No. 2,449,922, and the later art. The only apparent applicationinvolving the use of tubes under 1 mm O.D. is that of Christen et al,U.S. Pat. No. 4,098,852, which employs osmotic or ultrafilteringpolymeric tubes and vaporizing liquids. Christen's patent also utilizesthe shortest tubes found in the prior art in counterflow exchangers,such length being only about 0.6 m, compared to the more typical lengthof about 5 m. Roma's U.S. Pat. No. 4,030,540 is cited as a typicalexample of prior art design guidelines that often resulted in suchunsound objectives as attempting to maximize tube length, whereas thecorrect objective is always to minimize tube length while satisfyingseveral additional criteria.

Some useful related theoretical background materials may be found in twoof my earlier patents, although these inventions are quite remote fromthe present invention; U.S. Pat. No. 4,321,962 describes a solar energyheat exchanger and storage system; and U.S. Pat. No. 4,456,882 describesa high-speed turbine-driven air-bearing-supported sample spinner.

SUMMARY OF THE INVENTION

The present invention, the microtube-strip (MTS) counterflow heatexchanger, in the preferred embodiment consists of a number of heattransfer augmentation-free small modules connected in parallel. Eachmodule typically contains eight rows of one hundred tubes, each of 0.8mm outside diameter and 0.16 m length. The tubes are metallurgicallybonded to rectangular header tube strips at each end. Caps suitable formanifolding are welded over the ends. Means are provided to cause theshell-side fluid to flow in counterflow fashion over substantially allof the tube length, and suitable manifolds are provided to connect themodules in parallel. Power capacity per unit volume per unit temperaturedifference of the MTS exchanger exceeds that of prior art typicaldesigns by a factor of ten to 1000. Power capacity per unit cost perunit temperature difference of the MTS exchanger may exceed that ofprior art designs by a factor as large as 10 in some cases. Flowconditions in the microtubes are fully laminar and extremely subsonic.

Various other objects, features, and attendant advantages of the presentinvention will be more fully appreciated as the same becomes betterunderstood from the following detailed description, when considered inconnection with the accompanying drawings, wherein:

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an isometric drawing of an MTS sub-assembly.

FIG. 2 is a plane section view of an MTS header.

FIG. 3 is an isometric drawing of an MTS module.

FIG. 4 illustrates two reinforcement techniques for MTS modulesoperating with high tube-side pressure.

FIG. 5 is an isometric drawing of a plurality of MTS modules manifoldedtogether in parallel to form an MTS block.

FIG. 6 illustrates an MTS block enclosed in a pressurized vessel.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT The Heat Exchange Power

The usual approach to heat transfer problems is to begin with thefollowing equation:

    P.sub.h =h A T.sub.δ                                 (1)

where P_(h) is the heat transfer power (W), h is the heat transfercoefficient (W/m² K), A is the surface area (m²), and T.sub.δ is thetemperature difference (K). The problem then is to determine suitableexpression for h under various conditions. Unfortunately, mostengineers, after looking at equation (1), thereafter tacitly assume thatthe heat exchange power is proportional to the total surface area. It isthis erroneous underlying assumption that has virtually stagnatedprogress in signal phase heat exchanger design for four decades. Theoften overlooked fact is that the complicated heat transfer coefficient,h, is always inversely dependent on a characteristic dimension of theheat exchanger, often in such a way that P_(h) increases only as thesquare root of the area. In some case, P_(h) may be independent ofcertain changes in the area, and in other cases P_(h) may actually bedecreased by an increase in the area.

Consider first, for example, the tube-bundle heat exchanger with highturbulent gas flowing through the tubes, which are bathed in a constanttemperature fluid. The conventional approach is to write the heattransfer coefficient in terms of the dimensionless Nusselt number, Nu.

    Nu=hd/k,                                                   (2)

where d is the inside diameter (m) of the tubes and k is the thermalconductivity (Wm⁻¹ K⁻¹) of the gas. The Nusselt number is then expressedin terms of two additional dimensionless groups, the Prandtl number, Pr,and the Reynolds number, Re.

    Pr=C.sub.p μ/k,                                         (3)

where C_(p) is the constant pressure specific heat (J/kgK), and μ is thedynamic viscosity (kgm⁻¹ s⁻¹).

    Re=ρvd/μ=4G/πμd                               (4)

where ρ is the density of the gas (kg/m³), v is the mean velocity of thegas (m/s), and G is the mass flow rate per tube (kg/s). Then, for highlyturbulent flow it can be demonstrated that,

    Nu=0.023Pr.sup.4 Re.sup..8                                 (5)

Combining equations (2) through (5) gives the following expression forthe heat transfer coefficient.

    h=0.023(k/d)Pr.sup..4 (4G/πμd).sup..8                (6)

Thus, for a given turbulent mass flow rate through a bundle of tubes oflength L, the heat exchange power of equation (1) is proportional to thelength, and inversely proportional to the 0.8 power of the diameter.Hence, increasing the area by increasing the tube diameter actuallydecreases the heat exchange power, and the advantages of short tubes ofsmall diameter are readily apparent.

Now consider the case of a tube-type counterflow laminar-flow heatexchanger with center-to-center tube spacing equal to 1.4 times theoutside diameter of the tubes and twice the inside diameter. Furtherassume that the thermal conductivity of the tube material is muchgreater than the thermal conductivity of the fluids. For this case, itcan be shown that the heat exchange power is independent of the tubediameter, and is given by the following expression: ##EQU1## where n isthe number of tubes, k₁ is the thermal conductivity of the inner fluid,and k₂ is the thermal conductivity of the outer fluid.

From the above discussion it appears that there is little utility inevaluating a heat exchanger in terms of a heat exchange coefficient ofdimensions Wm⁻² K⁻¹ as is customary in the professional and patentliterature. Rather, a more useful characterization is the totaleffective flow length, nL. By defining nL as the quotient of P_(h) and ageneralized function of k₁ and k₂, one arrives at a useful method ofcomparing diverse designs-including those which incorporate heattransfer augmentation means such as extended or roughened surfaces.

Power Losses

The power, P_(p1) required to pump a fluid through the heat exchangertubes is given by:

    P.sub.p1 =(Δp)A.sub.f v,                             (8)

where Δp is the pressure drop (Pa) through the exchanger, A_(f) is thefrontal fluid area (m²), and v is the mean fluid velocity (m/s).

For simplicity, consider the case of laminar fluid flow through long,smooth tubes. This condition exists for Reynolds numbers, Re, below2000. The pressure drop, Δp, in a fluid flowing through a tube underlaminar conditions is given by:

    Δp=32μLv/d.sup.2,                                 (9)

thus:

    P.sub.p1 =8πμnLv.sup.2.                              (10)

The shell-side pumping power loss, P_(p2), required to pump fluid aroundthe tubes can be expressed by a similar equation:

    P.sub.p2 =fμnLv.sup.2,                                  (11)

where the gas parameters μ and v now refer to the external gas, and thecoefficient f is a complicated function of tube diameter and spacing.For the standard hexagonal-close-pack pattern with the distance betweentube centers equal to 1.4 times the tube outside diameter, f isapproximately equal to 200.

In addition to the pumping power loss, there is another internal lossmechanism present in counterflow exchangers which may limit thethermodynamic efficiency: the axial thermal conduction power of the tubemetal, P_(m).

    P.sub.m =πdnwk.sub.m (T.sub.H -T.sub.C)/L,              (12)

where w is the wall thickness of the tubes (m), k_(m) is the thermalconductivity of the tube metal (Wm⁻¹ K⁻¹), T_(H) is the mean temperatureat the hot end, and T_(C) is the mean temperature at the cold end.

Optimization

The power available, P_(i), from the input gas is:

    P.sub.i =GC.sub.p (T.sub.H -T.sub.C),                      (13)

where C_(p) is the constant pressure specific heat (J/kgK), and G is themass flow rate (kg/S) and is equal to ρA_(f) v. The waste heat, P_(o) is

    P.sub.o =GC.sub.p T.sub.δ,                           (14)

where T.sub.δ is, as defined earlier, the mean temperature differencebetween the counterflowing gases.

Accounting for the losses, the available heat exchange power, P_(E), is##EQU2## Equating input and output power gives, under steady-stateconditions, the following:

    P.sub.i +2P.sub.p1 =P.sub.E +P.sub.o,                      (16)

The above equations can now be solved for T₆₇ using the definition ofmass flow rate and assuming w=d/3. ##EQU3##

This equation depends only on three geometric variables, n, L, and d,and is reasonably valid for tube-type counterflow laminator heatexchangers, subject to several above mentioned assumptions. One can nowcalculate the power losses and the available heat exchange power for agiven set of thermodynamic and geometric conditions. The design can beoptimized via the linear programming technique of maximizing anobjective function, F_(c), such as the following:

    F.sub.c =(P.sub.E -aP.sub.p -bP.sub.o)/(total cost),       (18)

where a and b may have values of 10 and 2 respectively. It becomesapparent after exercising a linear programming technique on equation(18) that by giving proper attention to minimizing costs associated withtube cutting and end preparation, header hole punching, and tubeassembly and insertion techniques, optimized high power single phaseheat exchangers take on a totally new appearance. They consist ofhundreds or perhaps thousands of small modules, each of which consistsof hundreds of small, short tubes. Reynolds numbers inside themicrotubes for these optimized designs range from 25 to 400, compared tothe more common prior art values of 10,000 to 100,000; and Nusseltnumbers are less than 5, compared to the typical prior art values of 20to 400. The result is fully developed laminar flow, tube side and shellside, and flow velocities below one tenth the speed of sound.

Alternatively one may choose as objective function F_(v) such that

    F.sub.v =(P.sub.E -aP.sub.p -bP.sub.o)/(total volume),     (19)

Astoundingly, this function is unbounded. In other words, it istheoretically possible to increase the power-to-volume ratio withoutlimit, without increasing pumping losses, if one can reduce the tubediameter and length and increase the number of tubes without limit. Ofcourse, the above equations cease to be valid under molecular flowconditions.

The Tubing

Current practice in tube-type counterflow exchangers generally usesinduction-welded steel, copper, or aluminum tubes of about 3 mm to 25 mmdiameter with lengths ranging from 0.5 to 6 m and wall thickness ofabout 0.25 mm to 3 mm. However, recent advances in high speed laserwelding and super-hard die technology now make it possible to producevery small stainless steel hypodermic tubing at very low productioncosts-less than $0.10 per meter. It is thus practical to consider theuse of tubing with an outside diameter of less than 1 mm.

Reducing the tubing diameter by a factor of 10 requires the length to bereduced by a factor ranging from 30 to 100 while the number of tubes isincreased by a similar factor in order to maintain the same heatexchange power and pumping power loss. However, the total volume of theheat exchanger is likewise reduced. Furthermore, the maximum internalpressure rating of the heat exchanger will probably be increased due toan increase in the relative wall thickness.

To facilitate rapid assembly of large numbers of small tubes, it isnecessary to depart from the disc shaped tube header sheet normally usedin heat exchangers and instead use a rectangular tube header sheet orstrip. Furthermore, to minimize tube flexing and to reduce supportrequirements, it is also desirable to keep the tube length relativelyshort. This will also insure that the buckling strength of the tubes islarge enough to permit pressing them into the tube strip. Moreover, itwill raise the transverse acoustic resonance modes of the tubes therebymaking it more difficult to excite such resonances by turbulence. Also,equations 10 and 11 show that reducing the tube length will reduce thepumping power losses.

The maximum practical tube length for high-modulus, high strength alloyssuch as strain-hardened stainless steel or precipitation-hardenedsuperalloys is about 300 times the outside diameter of the tubes, whilethe maximum practical length for copper or aluminum tubes is about halfthat amount. There are several additional reasons for preferringstainless steel or superalloys over the more common heat exchangermetals: (1) They have very low thermal conductivity which may make themeasier to laser weld, but most importantly reduces the internal axialconduction loss mechanism, P_(m), in the counterflow exchangers; (2)Their high tensile strength allows higher working pressures; and (3)Their corrosion and high temperature strength properties are essentialin many applications.

Welding and Manifolding

The key to the current invention is the recognition of the advantage ofusing small diameter tubing in very short lengths. Its implementationdepends on technological breakthroughs in the assembly, welding, andmanifolding of these tubes. Since the tubes are very short, it isnecessary to resort to narrow modules in order that counterflowconditions be established over the major portion of the tube length andalso to reduce the inefficiencies due to non-uniform flow. While across-flow arrangement could be used to circumvent the above mentionednon-uniform flow problems, such as arrangement would greatly reduce thethermodynamic efficiency. The counterflow-serial-crossflow arrangementcommonly used in large installations allows somewhat higher efficiencythan the crossflow arrangement but at increased pumping losses. Hence,the most satisfactory solution is that of narrow modules of four totwenty rows of tubes.

The extremely small size of the tubes makes almost all types ofconventional welding methods impractical, and the extremely large numberof tubes eliminates most types of individual tube welding techniques,probably including automated electron beam and laser techniques becauseof process control problems arising from thermal expansion during thewelding operations. Two viable options for the tube-to-strip welds arefluxless brazing and diffusion welding. A wide variety of conventionalwelding techniques are suitable for the rest of the welds.

In the fluxless brazing technique, the braze metal is plated onto theinside of the holes and onto the outside of the tubes prior to assembly.After assembly, the complete module is heated in vacuum or inertatmosphere to the liquidus temperature of the braze metal. This methodis not suited for very high temperature exchangers.

Diffusion welding can be accomplished if the tube diameter and hole sizecan be held to very tight tolerances. The use of hardened tubes andannealed tube strips then makes it possible to press the tubes intoslightly undersized holes. With proper attention to surface quality anda minimum of 0.3% interference press fit, a strong metallurgical bondcan be formed simply by heating the assembly to about 0.8 times theabsolute melting temperature (K). This method is suitable for thehighest temperatures and all alloys.

Corrosion

In many cases, heat exchangers must operate in severely corrosiveenvironments. Under these conditions, it is no longer theoreticallypossible to increase the power-to-volume ratio without limit. Thecurrent state-of-the-art in corrosion resistant alloys, such as Nimonic81, limits the minimum wall thickness of about 50 microns for moderatelycorrosive environments and about 200 microns for severely corrosiveenvironments. Although the tubes themselves are too small to makecoatings or laminations practical with current technology, such measuresmay be applied to the tube strips and to the manifolds for economy ofmaterials or to achieve combined high temperature strength and hotcorrosion resistance.

Thermal Response Time

In many applications, particularly in the case of mobile gas turbines,fast response times are necessary for efficient operation. Currently, atypical 2000 KW gas turbine may have a mechanical response time of oneminute, but the thermal response time of the heat exchangersincorporated into the system may be ten hours. Increasing thepower-to-mass ratio of the heat exchanger by the amount possible withthe MTS design could reduce the thermal time constant to less than oneminute. Such a dramatic reduction in mass and thermal time constantopens up many new applications in all areas of transportation-especiallyaerospace.

High Pressure Applications

In many applications, for example, in recuperators used in closed cyclegas turbines, it is necessary to maintain both the internal (tube-side)and the external (shell-side) fluids at high pressure. The narrow widthof the tube header strip makes this design well suited to high tube-sidepressures. When high shell-side pressures are required, the entire heatexchanger must be enclosed in a pressurized containment vessel. Thesmall size of the heat exchanger simplifies this task.

DETAILED DESCRIPTION OF THE DRAWINGS

The basic unit in the MTS heat exchanger is the MTS sub-assembly asillustrated in FIG. 1. It consists of typically eight rows of heattransfer augmentation free microtubes 1 with typically 40 to 200microtubes in each row. The microtubes are diffusion welded intoprecision MTS header strips 2 at each end. The diffusion welding isaccomplished by using ultra precision, diamond-die-reduced, laser weldedhard drawn tubing for the microtubes, and precisely machining the holesin the annealed header strip to a size at least 0.3% smaller but notmore than 5% smaller than the tubing outside diameter. A combination oftechniques may be required to produce the precision holes in the headerstrips, including feinblanking, electrochemical machining, and reaming.The diffusion welds are accomplished by (1) insuring that the tubes andholes have thoroughly cleaned, oxide-free surface prior to assembly, (2)maintaining a minimum of 0.3% interference press fit, (3) heating thesub-assembly in an inert atmosphere or vacuum to a temperature ofapproximately 80% of the absolute melting temperature of the tube orheader strip alloy, whichever is lower.

FIG. 2 illustrates the recommended HCP (hexagonal close pack) holepattern for the MTS header strip 2. The distance between rows is equalto 0.866 times the distance between tube centers, TC, which is generallyabout 1.3 to 2.8 times the O.D. of the sample tubes 1.

FIG. 3 illustrates the basic counterflow MTS module. It includes asemi-cylindrical cap 3 welded to each header strip. Care is taken toassure that the header strip 2 is no wider than is necessary toaccommodate the microtubes 1 and the relatively thin walled cap 3 sothat the MTS modules may be mounted closely in parallel. Tube-sidemanifold ports 4 are provided on each cap 3. A cage 5 closely surroundsthe MTS sub-assembly, except near each header strip, forcing shell-sidefluid 6 to enter around the periphery of the MTS sub-assembly near oneend and to exit in like fashion at the other end. Tube-side fluid 7enters the tube-side manifold ports 4 at the end at which the shell-sidefluid exits, and it exits in like manner at the opposite end.

In certain applications, extremely high tube-side pressures, perhapscombined with very high temperatures, may require additional support ofthe flat header strip 2, to prevent bowing of this surface. Thisadditional support may be provided as shown in FIG. 4 by diffusionwelding a reinforcement plate 8 similar to the header strip 2 a shortdistance from it. Alternatively, the required support may be provided bythe microtubes 1 if they are supported in such a way to prevent theirbuckling. This may be accomplished by bonding, preferably by projectionwelding, stiffening wires 9 crosswise between the rows of microtubes 1.By staggering or offsetting the location of adjacent stiffening wires 9,the effect on fluid flow is generally made negligible.

FIG. 5 illustrates the parallel manifolding of several MTS modules toform an MTS block. Individual fluid ports 4 are connected to a tube-sidemanifold 10 at each end. The manifold cages 11 in cooperation with theMTS module cages 5 form the shell-side sealed region. Tube-size fluidmay exit at tube-side manifold port 12 while shell-side fluid may enterat manifold cage port 13. The MTS modules are supported by the headers,with adequate clearance space between the adjacent caps to permit therequired shell-side flow 6 between caps with acceptable pressure drop.Typical MTS blocks may include four to fifteen MTS modules in parallel,and typical high power installations may include hundreds of such MTSblocks further manifolding in parallel.

FIG. 6 depicts an MTS block mounted inside a pressure vessel 14 formingan MTS tank for applications requiring high shell-side pressures.Pressure equalizing vents 15 are required to equalize mean staticpressure components on the flat surface of the MTS cages 5 and manifoldcages 11. The dynamic pressure components arising from the shell-sidefluid pressure drop through the MTS block must be kept relatively smallto prevent excessive deflection of the flat surfaces. Expansion joints16 are required at one end to relieve axial thermal stresses. Suitablysealing flanges 17 and 18 are provided to permit convenient assembly ofthe containment vessel 14 and adequate sealing around the ports 12 and13. Suitable radial support for the MTS block within the vessel isrequired at the end which includes the expansion joints 16.

Although this invention has been described herein with reference tospecific embodiments, it will be recognized that changes andmodifications may be made without departing from the spirit of thepresent invention. All such modifications and changes are intended to beincluded within the scope of the following claims.

What is claimed is:
 1. A gas-gas laminar-flow heat exchanger modulewhich comprises:a plurality of heat transfer augmentation-free corrosionresistant, precision, hardened, metallic tubes .[.arrayed in at leastfour parallel disposed planar rows of at least forty tubes per row.]..Iadd.arrayed in a plurality of rows of tubes.Iaddend.; a firstrectangular header strip interference press fit and diffusion welded toone end of each of said tubes; a second rectangular header stripinterference press fit and diffusion welded to the other end of each ofsaid tubes; first manifold means metallurgically connected to said firstrectangular header strip for defining a gas inlet flow path into saidone end of each of said tubes; second manifold means metallurgicallyconnected to said second rectangular header strip for defining a gasoutlet flow path from said other end of each of said tubes; meansdisposed externally of said tubes for defining a counterflow flow-pathof heat exchanger gas over substantially the entire length of theexternal surfaces of each of said tubes from within the vicinity of saidother end of each of said tubes to within the vicinity of said one endof each of said tubes; each of said tubes having an outside diameter ofless than 3 mm; each of said tubes having a length which is sufficientto allow for fully developed laminar flow and which is less than 300times and outside diameter of each of said tubes; and said plurality oftubes within each of said rows being laterally spaced by acenter-to-center distance of from 1.3 to 2.8 times the outside diameterof each of said tubes.
 2. A module according to claim 1 wherein saidtubes have been produced from high tensile strength alloy metal.
 3. Amodule according to claim 2 wherein said high strength metal alloy isstainless steel.
 4. A module according to claim 1 wherein saidrectangular header strips are reinforced against high gas pressure byreinforced plates.
 5. A module according to claim 1 wherein said tubesare supported at one or more locations by stiffening wires weldedbetween said rows.
 6. A module according to claim 1 wherein said outsidediameter of each of said tubes is approximately 0.8 mm.
 7. A moduleaccording to claim 6 wherein said length of each of said tubes isapproximately 0.16 m.
 8. A module according to claim 1 wherein saidarray of tubes comprises .[.eight.]. .Iadd.at least 4 parallel disposedplanar .Iaddend.rows of said tubes .[.with from 40 to 200.]. .Iadd.withat least 40 .Iaddend.tubes per row.
 9. A module according to claim 1wherein said tubes are disposed within said rectangular header stripsthrough means of a 0.3% to 5% interference press fit.
 10. A moduleaccording to claim .[.1.]. .Iadd.8 .Iaddend.wherein the distance betweensaid parallel rows is equal to 0.866 times said center-to-center tubedistance.
 11. A module according to claim 1 wherein said means externalof said tubes for defining said counterflow flow-path of said heatexchanger gas comprises a cage annularly surrounding said array oftubes.
 12. A module according to claim 1 further comprising a pluralityof said modules defined by said tubes, said first and second headerstrips, and said first and second manifold means, vertically stackedtogether; and third and fourth manifold means, respectively connectingtogether said sets of first second manifold means of each of saidmodules, for supplying said tube-side gas inlet and outlet paths fromsaid modules; whereby said modules and said third and fourth manifoldmeans define a heat exchanger module block.
 13. A module according toclaim 12 further comprising pressure vessel means for housing said blockand thereby defining a heat exchanger module tank.
 14. A moduleaccording to claim 13 further comprising cage means surrounding andenclosing said first and second manifold sets and said third and fourthmanifold means for defining a counter-flow flow-path of heat exchangergas externally of said plurality of tubes.
 15. A module according toclaim 14 wherein said cage means are disposed internally within saidpressure vessel means.
 16. A module according to claim 15 furthercomprising expansion joint means defined between at least one of saidcage means and said pressure vessel means for relieving axial thermalstresses. .Iadd.17. A module according to claim 8 wherein said array oftubes comprises eight rows of said tubes with from 40 to 200 tubes perrow. .Iaddend.